The present invention pertains to improvements to an engine and more particularly to improvements relating to mechanical components of a Stirling cycle heat engine or refrigerator which contribute to increased engine operating efficiency and lifetime.
Stirling cycle machines, including engines and refrigerators, have a long technological heritage, described in detail in Walker, Stirling Engines, Oxford University Press (1980), herein incorporated by reference. The principle underlying the Stirling cycle engine is the mechanical realization of the Stirling thermodynamic cycle: isovolumetric heating of a gas within a cylinder, isothermal expansion of the gas (during which work is performed by driving a piston), isovolumetric cooling, and isothermal compression. The Stirling cycle refrigerator is also the mechanical realization of a thermodynamic cycle which approximates the ideal Stirling thermodynamic cycle. In an ideal Stirling thermodynamic cycle, the working fluid undergoes successive cycles of isovolumetric heating, isothermal expansion, isovolumetric cooling and isothermal compression. Practical realizations of the cycle, wherein the stages are neither isovolumetric nor isothermal, are within the scope of the present invention and may be referred to within the present description in the language of the ideal case without limitation of the scope of the invention as claimed. Various aspects of the present invention apply to both Stirling cycle engines and Stirling cycle refrigerators, which are referred to collectively as Stirling cycle machines in the present description and in any appended claims.
The principle of operation of a Stirling engine is readily described with reference to FIGS. 1a-1e, wherein identical numerals are used to identify the same or similar parts. Many mechanical layouts of Stirling cycle machines are known in the art, and the particular Stirling engine designated generally by numeral 10 is shown merely for illustrative purposes. In FIGS. 1a to 1d, piston 12 and a displacer 14 move in phased reciprocating motion within cylinders 16 which, in some embodiments of the Stirling engine, may be a single cylinder. Typically, a displacer 14 does not have a seal. However, a displacer 14 with a seal (commonly known as an expansion piston) may be used. Both a displacer without a seal or an expansion piston will work in a Stirling engine in an xe2x80x9cexpansionxe2x80x9d cylinder. A working fluid contained within cylinders 16 is constrained by seals from escaping around piston 12 and displacer 14. The working fluid is chosen for its thermodynamic properties, as discussed in the description below, and is typically helium at a pressure of several atmospheres. The position of displacer 14 governs whether the working fluid is in contact with hot interface 18 or cold interface 20, corresponding, respectively, to the interfaces at which heat is supplied to and extracted from the working fluid. The supply and extraction of heat is discussed in further detail below. The volume of working fluid governed by the position of the piston 12 is referred to as compression space 22.
During the first phase of the engine cycle, the starting condition of which is depicted in FIG. 1a, piston 12 compresses the fluid in compression space 22. The compression occurs at a substantially constant temperature because heat is extracted from the fluid to the ambient environment. In practice, a cooler (not shown) is provided. The condition of engine 10 after compression is depicted in FIG. 1b. During the second phase of the cycle, displacer 14 moves in the direction of cold interface 20, with the working fluid displaced from the region of cold interface 20 to the region of hot interface 18. This phase may be referred to as the transfer phase. At the end of the transfer phase, the fluid is at a higher pressure since the working fluid has been heated at constant volume. The increased pressure is depicted symbolically in FIG. 1c by the reading of pressure gauge 24.
During the third phase (the expansion stroke) of the engine cycle, the volume of compression space 22 increases as heat is drawn in from outside engine 10, thereby converting heat to work. In practice, heat is provided to the fluid by means of a heater (not shown). At the end of the expansion phase, compression space 22 is full of cold fluid, as depicted in FIG. 1d. During the fourth phase of the engine cycle, fluid is transferred from the region of hot interface 18 to the region of cold interface 20 by motion of displacer 14 in the opposing sense. At the end of this second transfer phase, the fluid fills compression space 22 and cold interface 20, as depicted in FIG. 1a, and is ready for a repetition of the compression phase. The Stirling cycle is depicted in a P-V (pressure-volume) diagram as shown in FIG. 1e. 
Additionally, on passing from the region of hot interface 18 to the region of cold interface 20, the fluid may pass through a regenerator (not shown). The regenerator may be a matrix of material having a large ratio of surface area to volume which serves to absorb heat from the fluid when it enters hot from the region of hot interface 18 and to heat the fluid when it passes from the region of cold interface 20.
The principle of operation of a Stirling cycle refrigerator can also be described with reference to FIGS. 1a-1e, wherein identical numerals are used to identify the same or similar parts. The differences between the engine described above and a Stirling machine employed as a refrigerator are that compression volume 22 is typically in thermal communication with ambient temperature and expansion volume 24 is connected to an external cooling load (not shown). Refrigerator operation requires net work input.
Stirling cycle engines have not generally been used in practical applications, and Stirling cycle refrigerators have been limited to the specialty field of cryogenics, due to several daunting engineering challenges to their development. These involve such practical considerations as efficiency, vibration, lifetime, and cost. The instant invention addresses these considerations.
A major problem encountered in the design of certain engines, including the compact Stirling engine, is that of the friction generated by a sliding piston resulting from misalignment of the piston in the cylinder and lateral forces exerted on the piston by the linkage of the piston to a rotating crankshaft. In a typical prior art piston-crankshaft configuration such as that depicted in FIG. 2, a piston 10 executes reciprocating motion along longitudinal direction 12 within cylinder 14. Piston 10 is coupled to an end of connecting rod 16 at a pivot such as a pin 18. The other end 20 of connecting rod 16 is coupled to a crankshaft 22 at a fixed distance 24 from the axis of rotation 26 of the crankshaft. As crankshaft 22 rotates about the axis of rotation 26, the connecting rod end 20 connected to the crankshaft traces a circular path while the connecting rod end 28 connected to the piston 10 traces a linear path 30. The connecting rod angle 32, defined by the connecting rod longitudinal axis 34 and the axis 30 of the piston, will vary as the crankshaft rotates. The maximum connecting rod angle will depend on the connecting rod offset on the crankshaft and on the length of the connecting rod. The force transmitted by the connecting rod may be decomposed into a longitudinal component 38 and a lateral component 40, each acting through pin 18 on piston 10. Minimizing the maximum connecting rod angle 32 will decrease the lateral forces 40 on the piston and thereby reduce friction and increase the mechanical efficiency of the engine. The maximum connecting rod angle can be minimized by decreasing the connecting rod offset 24 on the crankshaft 22 or by increasing the connecting rod length. However, decreasing the connecting rod offset on the crankshaft will decrease the stroke length of the piston and result in less xcex94 (pV) work per piston cycle. Increasing the connecting rod length can not reduce the connecting rod angle to zero but does increase the size of the crankcase resulting in a less portable and compact engine.
Referring now to the prior art engine configuration of FIG. 3, it is known that in order to reduce the lateral forces on the piston, a guide link 42 may be used as a guidance system to take up lateral forces while keeping the motion of piston 10 constrained to linear motion. In a guide link design, the connecting rod 16 is replaced by the combination of guide link 42 and a connecting rod 16. Guide link 42 is aligned with the wall 44 of piston cylinder 14 and is constrained to follow linear motion by two sets of rollers or guides, forward rollers 46 and rear rollers 48. The end 50 of guide link 42 is connected to connecting rod 16 which is, in turn, connected to crankshaft 22 at a distance offset from the rotational axis 26 of the crankshaft. Guide link 42 acts as an extension of piston 10 and the lateral forces on the piston that would normally be transmitted to cylinder walls 44 are instead taken up by the two sets of rollers 46 and 48. Both sets of rollers 46 and 48 are required to maintain the alignment of guide link 42 and to take up the lateral forces being transmitted to the guide link by the connecting rod. The distance d between the forward set of rollers and the rear set of rollers may be reduced to decrease the size of the crankcase (not shown). However, reducing the distance between the rollers will increase the lateral load 54 on the forward set of rollers since the rear roller set acts as a fulcrum 56 to a lever 58 defined by the connection point 52 of the guide link and connecting rod 16.
The guide link will generally increase the size of the crankcase because the guide link must be of sufficient length that when the piston is at its maximum extension into the piston cylinder, the guide link extends beyond the piston cylinder so that the two sets of rollers maintain contact and alignment with the guide link.
In accordance with one aspect of the invention, a system for supporting lateral loads on a piston undergoing reciprocating motion along a longitudinal axis in a cylinder includes a guide link coupling the piston to a crankshaft undergoing rotary motion about a rotation axis of the crankshaft. A first guide element is located along the length of the guide link and includes a spring mechanism for urging the first guide element into contact with the guide link. The spring mechanism includes a first spring with a first natural frequency of oscillation and a second spring with a second natural frequency of oscillation. A second guide element is in opposition to the first guide element. In one embodiment, the first guide element is a roller having a rim in rolling contact with the guide link and the second guide element is a roller with a rim in rolling contact with the guide link.
In a further embodiment, the second guide element includes a precision positioner for positioning the second guide element with respect to the longitudinal axis. The precision positioner may be a vernier mechanism having an eccentric shaft for varying a distance between the second guide element and the longitudinal axis.
In accordance with another aspect of the invention, a linkage for coupling a piston undergoing reciprocating linear motion along a longitudinal axis to a crankshaft undergoing rotary motion about a rotation axis of the crankshaft includes a guide link having a first end proximal to the piston and coupled to the piston and a second end distal to the piston such that the rotation axis is disposed between the proximal end and the distal end of the guide link. A connecting rod is rotably connected to the end of the guide link distal to the piston at a rod connection point at a connecting end of the connecting rod. The connecting rod is coupled to the crankshaft at a crankshaft connection point on a crankshaft end of the connecting rod, where the crankshaft connection point is offset from the rotation axis of the crankshaft. A guide link guide assembly supports lateral loads at the distal end of the guide link and includes a first roller having a center of rotation fixed with respect to the rotation axis of the crankshaft and a rim in rolling contact with the distal end of the guide link. A spring mechanism is used to urge the rim of the first roller into contact with the distal end of the guide link. The spring mechanism includes a first spring with a first natural frequency of oscillation and a second spring with a second natural frequency of oscillation.
In one embodiment, the guide link guide assembly further includes a second roller in opposition to the first roller and having a center of rotation and a rim in rolling contact with the distal end of the piston. The second roller may include a precision positioner to position the center of rotation of the second roller with respect to the longitudinal axis. In a further embodiment, the precision positioner is a vernier mechanism having an eccentric shaft for varying the distance between the center of rotation of the second roller and the longitudinal axis.
In accordance with yet another aspect of the invention, an improvement is provided to a Stirling cycle machine of the type where at least one piston undergoes reciprocating motion along a longitudinal axis in a cylinder. The piston is coupled to a crankshaft undergoing rotary motion about a rotation axis using a guide link having a first end proximal to the piston and coupled to the piston and a second end distal to the piston. The improvement has a guide link guide assembly including a spring mechanism for urging the rim of a first roller into contact with the distal end of the guide link where the spring mechanism includes a first spring with a first natural frequency of oscillation and a second spring with a second natural frequency of oscillation.